Cushion stop for hydraulic actuators



13, 1968 M. c. DARLING 3,396,635

CUSHION STOP FOR HYDRAULIC ACTUATORS Filed Nov. 30, 1966 3 Sheets-Sheet l INVENTOR. F I yARv/N a DARLl/VG A T TOR Aug. 13, 1968 M. c. DARLING CUSHION STOP FOR HYDRAULIC ACTUATORS 5 Sheets-Sheet 2 Filed Nov 30, 1966 PRIOR ART FIG.

MAN

MAN

46? HYD HYD 47 INVENTOR. MARV/N C. DARLl/VG BY M FIG.

A 7' TOR/V5 Y Aug. 13, 1968 I M. c. DARLING 3,396,635

CUSHION STOP FOR HYDRAULIC ACTUATORS Filed Nov. 30, 1966 3 Sheets-Sheet 3 FIG.4

INVENTOR. MAR V/lV C. 0A RLl/VG fifim/m ATTORNEYS United States Patent 3 396,635 CUSHIGN STOP FOKHYDRAULIC ACTUATORS Marvin C. Darling, Burrton, Kans., assignor to The Cessna Aircraft Company, Wichita, Kans., a corporation of Kansas Filed Nov. 30, 1966, Ser. No. 598,040 9 Claims. (Cl. 91-396) ABSTRACT OF THE DISCLOSURE The invention is an improved cushioning stop control for hydraulic motors moving large inertia loads. Upon approaching the end of the stroke the fluid being discharged from the exhaust chamber of the motor is progressively restricted, causing an increased discharge pressure which actuates a by-pass valve to divert fluid pressure from the inlet chamber of the motor to a reservoir, thereby decreasing the overall amount of energy necessary to bring the system to a rest.

The invention relates to a cushioned stop device for use in both linear and rotary hydraulic motors which utilize some type of progressively restrictive discharge orifice at the end of the stroke.

More particularly, the invention is directed to a means for a cushioned stop without the damaging high pressure buildup in the exhaust chamber of a motor. The snubbing condition is created by reducing the pressure in the inlet chamber at the same time the pressure in the exhaust chamber begins to increase. The control mechanism, upon sensing an increased discharge pressure by reason of the restrictive orificing, diverts fluid from the inlet chamber to a low pressure reservoir. The effect of the inlet pressure drop provides a comparable cushion stop with lower exhaust chamber pressures. The control mechanism which also functions as a piston velocity governor, has particular utility in actuators which move large inertia loads.

Fluid motors of the present day are expected to complete a stroke as quickly as practical and within the limits of economical hydraulics. Such a requirement necessitates hydraulic actuators which produce extremely high torque and speed capabilities. When connected to a large inertia load mass, such as a loaded crane boom, an excessive amount of energy may be transferred to the load mass during actuation. The absorption of such energies at the end of the stroke can cause dangerous high pressure buildups and shocks on the hydraulic system. Pressures of this order produce tremendous stresses and wear on piston heads, piston rods and other related mechanical linkages.

In the prior art, cushioning has been accomplished by progressively reducing the discharge orifice as the piston or vane approaches the end of its stroke. This progressive restriction of flow causes a pressure buildup in the discharge chamber by reason of the combined inertia effect of the load and the pressure in the inlet chamber. Although cushioning devices of this type are satisfactory for low performance actuators, they are a severe limitation on high speed actuators for use on heavy loads.

In the instant invention, the high exhaust chamber pressures are avoided by relieving the input pressure during the cushioning portion of the stroke. This allows a cushioned stop of a motor having a substantial inertial load with approximately one-half the pressure buildup in the exhaust chamber as in the above-mentioned prior art devices.

It is, therefore, the principal objective of the present invention to provide a new and improved cushioning means to bring the piston or vane of a hydraulic motor to a smooth stop at the end of its stroke with a minimal exhaust pressure.

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Another object of the invention is to provide a cushioning device for stopping an inertial loaded piston or vane by reducing the pressure in the inlet chamber during the cushioning portion of the stroke.

An additional object of the invention is to provide a hydraulic motor with an automatic control of the input pressure in a manner that the input pressure is reduced when the load on the motor tends to accelerate the piston or vane travel beyond a desired velocity.

A further object of the invention is to provide a hydraulic motor having a cushioning means whereby a pressure rise in the exhaust chamber actuates a bypass valve opening the input chamber to an exhaust line, thereby dropping the inlet pressure below the exhaust pressure.

Another object of the invention is to provide a cushion stop means which vents excessive pressures in the exhaust chamber by way of a high pressure relief valve to the inlet chamber.

A still further object of the invention is to provide a cushion stop in a hydraulic actuator having low pressure fluid supplied to the inlet chambers at all times to prevent a possible cavitating condition.

Further objects and advantages of the invention will be in part apparent and in part pointed out specifically hereinafter in connection with the description of the drawings that follows and in which:

FIGURE 1 is a combined graphical and pictorial view of a hydraulic cylinder, its associated power source, directional valves and automatic cushioning mechanism, all of which is shown partially cut away to illustrate the fluid flow;

FIGURE 2 is a diagrammatic view of a double acting lineal cylinder and the commonly known progressive orificing-type cushion stop;

FIGURE 3 is a diagrammatic view of the automatic cushioning mechanism of the present invention with a rotary cylinder; and

FIGURE 4 is a diagrammatic view of an automatic cushioning mechanism illustrating a variation which includes an anti-cavitation circuit.

Referring now to the drawings for a detailed description of the invention and, more specifically, to FIGURE 2, a commonly known method of cushioning hydraulic cylinders in the prior art is shown in diagrammatic form and is generally identified by reference numeral 10. The double acting lineal cylinder 11 is illustrated with a piston having a pair of piston rods 13 attached thereto. Connected to the outer ends of the rods 13 are Work loads 14 shown in symbolic form. The cylinder 11 derives its hydraulic energy from a constant volume pump 15 or the like which bypasses fluid when necessary through relief valve 16 to reservoir 18. The movement of the cylinder 11 is controlled by an ordinary directional control valve 19 which supplies fluid to the respective ends of the cylinder 11 through openings 20 and 21. Axially aligned with the discharge and inlet openings 20 and 21 are a pair of tapered pins 22 integral with the piston 12 which, along with said openings, provide a variable discharge orifice approximate the end of the stroke. For example, if chamber 23 were the inlet chamber, valve 19 would be in the straight through position allowing fluid to ingress through opening 20, thereby causing the piston to move to the right while fluid freely discharges through opening 21. As the piston 12 nears the end of its stroke, the orifice pin 22 will enter the discharge opening 21 and, because of its tapered design, will gradually meter the flow through an orifice of gradually decreasing size until the opening is completely closed. As the orifice 21 begins to close, the pressure in the discharge chamber 24 increases as it absorbs the energy in the system. If the work loads 14 are substantial, the inertia effect of pulling the cylinder combined with inlet pressure will produce momentary pressure in the discharge chamber 24 that can seriously damage the system as previously mentioned. Although FIGURE 2 illustrates a specific pin-type progressive orificing system, there are a variety of other types of variable orifices, some of which will be later described. In all types of variable orifices, the cushioning is produced by restricting the discharge flow, thereby severely lirniting the application of such cushioning devices.

The present invention, which is an improved cushion circuit, is diagrammatically shown in FIGURE 3 and generally identified by numeral 25. The cushioned stop circuit 25 is shown with a rotary cylinder 26 for illustrative purposes only. The progressive orificing structure shown in cylinder 26 is also merely one of many configurations which can be used.

The rotary cylinder 26 is schematically shown having its various ports entering through the end plate 27 of the cylinder. Integral with the load carry shaft 28 is the movable vane 29 which maintains sealing contact with the inner surface of the cylinder 26 during rotation. The fixed vane 30 integrally attached to the cylinder 26 also maintains sealing contact with the surface of the rotating shaft 28, thereby providing inlet and discharge chambers 31 and 32, respectively. The flow into the cylinder 26 is from line 33 through a series of small orifices 34 into inlet chamber 31 which causes the vane 29 to move in a clockwise direction. As the vane 29 approaches the end of its stroke, as shown in the drawing, the discharge orifices 35 are progressively closed by a spring biased plate 36 attached to and moving with the vane 29. The plate 36, although attached to the vane 29, is free to move up oil? of the surface of the end plate 27. When the flow is reversed, the light spring holding the plate 36 in contact with the end plate is lifted upward to permit full flow through the orifices 35. The angular velocity of the vane 29 and its respective inertia load are limited by the fixed restricting orifices 37 and 38 in lines 33 and 39. The directional control valve 40, pump 41, bypass valve 42 and reservoir 43 are all similar in operation to those shown in FIGURE 2 and, therefore, will not be described in detail. The two position automatic valves 44 and 45 will be later described in detail in the FIGURE 1 description. For purposes of FIGURE 3, valves 44 and 45 are actuated by pilot cylinders 46 and 47 which sense inlet and discharge pressures through openings 49 and 50. When the pilot cylinders are not extended, the valves 44 and 45 are spring biased to a closed position as seen in FIGURE 3.

Operation of the cushion stop with a rotary cylinder (FIGURE 3) When the control valve is moved to the straight through position, fluid is introduced into chamber 32. As the vane 29 begins to move, fluid is discharged from chamber 31 through orifices 34. The Velocity of the vane and its respective load is limited by the fixed restricting orifices 37 and 38. As the vane 29 and its slide valve plate 36 begin to progressively close the discharge openings 34, pressure builds in chamber 31. Said pressure is sensed by pilot cylinder 47 through line 54 and opening 49 to hydraulically actuate valve 45. With valve open, fluid is allowed to flow unrestricted out of chamber 32 through opening 50, bypass line 55, across check valve 57 and into return line 33. Since the flow into chamber 32 is restricted by orifice 38, pressure in said chamber drops to zero, thus creating a negative pressure differential across the vane 29. As the inertia load carried by the rotary shaft 28 is brought to a stop, the pressure drop in chamber 32 has a negative effect on the critical pressure surges in the discharge chamber 31. In the prior art cushion stops, such as mentioned above in FIGURE 2, the inlet pressure increases the discharge pressure while the negative pressure condition of the instant invention lessens the discharge chamber pressure required to stop a similar inertia force. The cushion stop of the present invention thereby permits the same fluid motor used in the prior art devices to handle greater loads at higher speeds without damage to the hydraulic system.

The corresponding automatic valve 44, bypass line 56, check valve 58 and opening 50 operate in the same manner as described above when the vane 29 rotates in the opposite direction and, therefore, are not herein described in detail. After the vane 29 comes to rest and the system becomes static, the automatic valve 45 is spring returned to its flow-interrupting position.

FIGURE 1 In this figure, the cushion stop structure is illustrated with a lineal hydraulic motor 60. The automatic valve mechanisms 44 and 45 described above in schematic form are shown in detail in this figure. Any type of lineal or rotary fluid motor can be used with the novel cushioning structure of the present invention as can be seen in the drawings.

Although the cylinder 60 is shown separate from the valve mechanisms 44 and 45, they can be incorporated together as a unitary structure. For example, in a rotary motor, the valve and porting mechanisms may be located in the fixed vane or in the rotary shaft.

The hydraulic power supply 62, which includes the pump 15, relief valve 16, reservoir 18 and the control valve 40, is similar in operation to that shown and described in FIGURE 2. The cylinder 60 is a double acting type having a piston 63 and piston rod 64 connected to a working load 65. Integral to the piston 63 are a pair of tapered orificing pins 66 and 67 having corresponding ports 68 and 69 in the end walls of the cylinder 60. The orificing pins and their respective ports progressively restrict the discharge flow as the end of the stroke is approached in the same manner as previously described.

The automatic valve mechanisms 44 and 45, which are normally located in the body of the fluid motor, are depicted separate therefrom for purposse of illustration and simplicity. The valve 44 is longitudinally positioned in bore 70. The valve 44 consists of a spool 73 slidably mounted in a sleeve 74. The spool 73 is normally held in the position shown in the left valve 44, by spring 75 pushing downward against spring retainer 76. The spring tension is adjusted by movement of the adjusting ring 77. The spool 73 may be moved upward, depressing the spring 75, by hydraulic pressure on piston 78 which will later be described in detail. The upward travel of the spool 73 is limited by tube 79 which in turn is stopped by a threaded port fitting 81. The tube 79 has openings 82 in its sides to allow fluid to pass from the tube 79 to the bore 70.

The fluid fiow through the valve mechanism 44 has two paths, the first of which is from port 83, through the lower end of tube 79, around the spring 75, past the annular flow restriction 84 (provided between the bore 70 and retainer 76), through orifice 85 and through line 86 into the upper chamber 83. The second fluid path is open when the spool 73 is in its upper position as illustrated in valve 45. The path is through bypass line 87, through aligned openings 89 and 90 and up the center passage 91 to join the flow of said first path.

With directional valve 40 in the straight through posi tion, fluid flows out line 94, through valve 44, as indicated by the arrows in the drawing, and into the piston chamber 88. As the piston 63 responds, fluid in the exhaust chamber 99 flows as indicated out line 92, through valve 45 to the return line 93. As the orificing pin 67 passes through port 69, the flow decreases and a pressure rise occurs in exhaust chamber 99. The increased pressure is sensed in line 95 causing the pilot piston 78 to exert an increased pressure on spool 73. The combined forces exerted by the piston 78 and the force caused at the restriction 84 by reason of the pressure drop are suflicient to override the spring 75 and raise the spool 73 to its upper limit, as seen in the right hand valve 45. The spool 73 in this position opens bypass line 87 through now aligned openings 89 and 90 so that fluid entering chamber 88 is bypassed to return line 93. By reason of the fixed restriction 85 in line 86, the pressure in chamber 88 drops to zero as the bypass line 87 is opened. Fixed restrictions 85 in lines 86 and 87 also function to limit the maximum velocity of piston 63. While the bypass lines 87 and 97 are essentially unrestricted, the inlet and discharge lines 86 and 92 have a series of restrictions therein. The discharge flow, for example, passes through orifices 69, 85 and 84, each of which have been described above.

When an abnormal inertia load 65 is experienced by the system, an additional energy absorption circuit comes into play. When a maximum pressure is reached in chamber 99, relief valve 100 opens and fluid flows from the discharge chamber 99 to inlet chamber 88 by way of line 104 through check valve 103. By reason of the above relief circuit, the maximum internal pressure in the cylinder can be controlled and, at the same time, can be used With a maximum efficiency to absorb kinetic energy acting on the cylinder 60. Relief valve 102 and check valve 105 in line 106 operate in the same manner as in the circuit just described when the piston 63 is being cushioned in the opposite direction.

Located in the bypass line is a ball check 107 which seats in the end of the spring retainer 76, blocking reverse flow through passage 91 of spool 73. The purpose of check 107 is to prevent reverse flow in the bypass line 87. Such a condition would only arise if control valve 40 reversed flow during the cushioned portion of the stroke (when spool 73 is extended as seen in the right hand valve 45). Under such a condition, the absence of check 107 would permit fluid entering line 93 to flow unrestricted through bypass line 87 to chamber 88. In addition, there is flow from chamber 99 out line 92 and around spring retainer 76 into passage 91 Where it joins the above-mentioned flow from the pump 15 into chamber 88. This condition allows a large amount of oil into chamber 88 which creates considerable back pressure in chamber 88 by reason of flow restrictions 85 and 84 in the left hand valve. This pressure will be added to the inertia load 65 already on piston 63. Such a condition would produce dangerously high pressures in chamber 99, which is contrary to the purpose of the present invention. With check 107, this reverse flow is prevented when the ball seats in the spring retainer 76 (as it ap pears in the left hand valve 44).

If the control valve 40 is switched from the straight through position to the criss-cross position while the piston is in mid-stroke (not in the cushion phase), the following will take place. The discharging flow from chamber 99 will meet the reversed pump flow entering valve 45 through line 93. The inertia eflect caused by the piston load 65 will cause a rapid pressure rise in chamber 99. Ball check 107 being closed, flow through bypass line 87 is blocked. As the pressure in chamber 99 reaches the maximum permitted, relief valve 100 opens permitting flow into chamber 88. Once the inertia force is overcome by the pressure differential across the piston 63 and the relief valve resistance, the piston will begin to move in the opposite direction under normal operating pressures as the relief valve 100 closes.

The lower end of spool 73 has a small passage 109 which permits fluid to enter chamber 110 as the spool 73 is being extended. A notch 112 is provided in the end of piston 78 which prevents the piston from blocking flow through passage 109. The presence of passage 109 provides a fiow restriction which controls the rate of travel of spool 73.

Automatic control of the input pressure, when the external load 65 tends to accelerate the piston velocity beyond the desired rate, is accomplished by the valve 45 in the following manner. When piston 63 is in mid-stroke and the load 65 attempts to accelerate the piston movement, back pressure builds up by reason of restrictions 69, and 84 in the discharge flow path. Sufiicient back pressure in chamber 99 will cause spool 73 to move upward, opening bypass line 87 and relieving the input pressure in chamber 88. The proper balance of areas between restrictions 69, 85 and 84, in addition to the proper compression setting of spring 75, will cause the input pressure to drop when the piston 63 reaches a maximum desired velocity, and also cause it to increase if the piston velocity drops below a desired rate.

FIGURE 4 EMBODIMENT An additional embodiment of the previously disclosed cushioning structure having a different anti-cavitation circuit is shown in this figure. The power source 62, including pump 15, relief valve 16, reservoir 18 and directional control valve 40, operates in the same manner as above-described. The cylinder 60, although shown in schematic form, is similar in structure and operation as that shown in FIGURE 1. Automatic valves 44 and 45 and their related lines, along with relief valves 100 and 102, are also the same as previously described.

The anti-cavitation circuit includes the relief valves 100 and 102, check valves 103 and 105, and lines 116, 117, 118, 120 and 122. If, at any time, excessive pressure is built up in chamber 99 and passage so that relief valve 100 opens, fluid from passage 95 passes through line 118, across relief valve 100 to line 117, over check valve to line 96. Excess pressure in chamber 99 is thus relieved by fluid flow to chamber 88. EX- cess pressure in chamber 88 is relieved in a similar manner.

Line acts as another anti-cavitation feature to that just described. Line 120 connects into low pressure return line 122 which dumps into reservoir 18. A small line pressure of approximately 50 psi. is maintained in line 122 and is schematically represented by restriction 124. Line 120 also connects with line 117. If pressure becomes negative in either chamber 88 or 99, fluid will flow from return line 120, across the appropriate check valve 103 or 105 into the chamber 88 or 99 to prevent possible cavitation.

Operatiom-FIGURE 4 embodiment If control valve 40 is suddenly closed with the piston 63 at mid-stroke, the inertia load 65 will cause the piston 63 to continue movement. A cavitation situation will arise in chamber 88 and high pressure will develop in chamber 99. Relief valve 100 will open, allowing fluid to flow to chamber 88, thereby relieving the threat of cavitation until the inertia force is overcome.

If a condition arises Where the control valve 40 remains open and there is a loss of pump pressure, a cavitation situation can occur by reason of the inertia effect of the load 65 on the piston 63. However, the small back pressure maintained in lines 120 and 96 is sufliciently great to supply adequate fluid to chamber 88 to avoid cavitation.

It is understood that variations from the form of this invention disclosed herein may be made without departure from the spirit and scope of the invention and that the drawings and specification are to be considered as merely illustrative.

What is claimed is:

1. In a hydraulic actuator construction for use with heavy inertia loads including a housing defining a Working chamber having a differential pressure movable barrier traversing said chamber and dividing it into two variable volume chambers both of which alternately act as an intake chamber having a port therein connected by a supply line to a source of hydraulic pressure while the other acts as a discharge chamber having a port connected to an exhaust line, an actuator shaft extending through the housing connecting the movable barrier to an exterior movable mass for transmitting energy from the fluids in the working chamber to said mass, variable orificing means cooperating with the discharge port for severely restricting the outflow of fluid therethrough near the end of the stroke of the movable barrier, the improvement comprising:

duct means independent of the supply line alfording communication between the intake chamber and an exhaust line;

valve means positioned in said duct means normally spring biased to a closed position whereby flow through the duct is blocked;

flow restriction means in the supply line restricting flow to the intake chamber to a rate less than the flow capacity of said duct means;

valve operator means responsive to a predetermined pressure in the discharge chamber, operatively connected to the valve means, to override the valve spring and maintain the valve in an open position as long as the predetermined pressure level is exceeded or maintained in the discharge chamber, whereby the pressure in the intake chamber is reduced substantially to zero as the barrier approaches the end of its stroke.

2. In a hydraulic actuator construction as set forth in claim 1, including:

passage means connecting the discharge chamber with the intake chamber;

a check valve positioned in said passage permitting flow only from the discharge chamber to the intake chamber;

a high pressure relief valve positioned in said passage permitting flow in the passage only when the pressure in the discharge chamber exceeds a preset pressure which is greater than the pressure needed to open said valve means.

3. In a hydraulic actuator construction as set forth in claim 1, including:

a first passage means connecting the discharge chamber with the intake chamber;

a check valve positioned in the first passage means permitting flow only into the intake chamber;

a high pressure relief valve positioned in the first passage between the discharge chamber and the check valve permitting flow in the first passage only when the pressure in the discharge chamber exceeds a preset pressure which is greater than the pressure needed to open said valve means;

a low pressure fluid source;

a second passage connecting the low pressure fluid source to the first passage at a point between the check valve and the relief valve.

4. In a hydraulic actuator construction as set forth in claim 1, including:

passage means connecting the discharge chamber with the intake chamber;

a check valve positioned in said passage permitting flow only from the discharge chamber to the intake chamber;

a high pressure relief valve positioned in said passage permitting flow in the passage only when the pressure in the discharge chamber exceeds a preset pressure which is greater than the pressure needed to open said valve means;

a second check valve positioned in said duct means preventing flow into the intake chamber.

5. In a hydraulic actuator construction as set forth in claim 1, including:

passage means connecting the discharge chamber with the intake chamber;

a check Valve positioned in said passage permitting flow only from the discharge chamber to the intake chamber;

a high pressure relief valve positioned in said passage permitting flow in the passage only when the pressure in the discharge chamber exceeds a preset pressure which is greater than the pressure needed to open said valve means;

fixed restriction means positioned in the ports of said intake and discharge chambers whereby the speed of said movable barrier is governed.

6. In a hydraulic actuator construction as set forth in claim 1 wherein:

the actuator is of a rotary type having a cylindrical Working chamber limited by planar end walls;

the movable barrier being a rotary piston journaled in the end walls having at least one vane in slidable sealing engagement with the cylindrical wall of the working chamber and the end walls thereof;

at least one fixed vane abutment radially positioned within the working chamber and fixed from movement with respect to said working chamber having slidable sealing engagement with said piston thereby providing two variable volume chambers separated by said fixed and movable vanes.

7. In a hydraulic actuator construction as set forth in claim 6, wherein the variable orificing means includes:

a series of discharge ports in one of the end walls located in close proximity to the fixed vane and positioned to be progressively closed by portions of the movable vane acting as a slidable valve face as the vane approaches the end of the stroke.

8. In a hydraulic actuator construction as set forth in claim 1, wherein:

the actuator is a lineal type cylinder having a cylindrical working chamber with planar end walls;

the movable barrier being a piston normally disposed in the working chamber for longitudinal movement therein;

the actuator shaft comprising a piston rod axially disposed in said working chamber connecting to said piston and longitudinally extending through at least one end wall thereof.

9. In a hydraulic actuator construction as set forth in claim 8, wherein the variable orificing means includes:

a tapered metering pin attached to the piston in axial alignment with the discharge port whereby as the pin penetrates into the port, the cross sectional area of the port is gradually decreased to Zero as the piston reaches the end of its stroke.

References Cited UNITED STATES PATENTS 2,028,766 1/1936 Ernst et al. 91-421 2,464,283 3/ 1949 Adams 91-421 2,493,602 1/ 1950 Sterrett 91-396 2,741,895 4/1956 Horvath 91-436 2,935,047 5/1960 Ortman et a1. 91-396 2,681,581 6/1954 Pearson 91-448 PAUL E. MASLOUSKY, Primary Examiner. 

